Improvements in rotary claw pumps

ABSTRACT

A Rotary Claw Pump with at least one pair of intermeshing rotors has the same shape within a tolerance of ±3% of the radii from edges to rotational centers. A convex portion of the claws has a nominal uniform rate of radius change with angle of rotation. There may be multiple stages wherein the thickness of each successive rotor pair maintains profile but has a thickness within 20% of the thickness of the first stage rotor pair multiplied by the inverse products of the compression ratios of the preceding stages. A discharge port is at least partially uncovered at any position of rotation from where intermeshing claw rotor tips are closest to where they are most distant. The discharge port is at least partially connected with a region connected with the rotor tips of both claws during rotation. The discharge port limits pressure increase in the region.

COPYRIGHT

A portion of the disclosure of this patent document contains material which is subject to copyright protection. The copyright owner has no objection to the facsimile reproduction by anyone of the patent document or the patent disclosure, as it appears in Patent and Trademark Office patent files or records, but otherwise reserves all copyright rights whatsoever.

TECHNICAL FIELD

The invention relates to the field of Pumps.

CITATIONS

-   U.S. Pat. No. 9,444,378 Compressed Air Energy Storage System     Utilizing Two-Phase Flow to Facilitate Heat Exchange. September 2016 -   U.S. Pat. No. 8,887,593 Device of a Pair of Claw-Type Rotors Having     Same Profiles. Nov. 18, 2014. -   U.S. Pat. No. 8,517,701 Multistage Vacuum Pump. Aug. 27, 2013. -   U.S. Pat. No. 8,308,458 Rotor Assembly for Multi-Stage Pump. Nov.     13, 2014. -   U.S. Pat. No. 7,594,323 Methods for Designing Single-Lobe and     Double-Lobe Rotors. Sep. 29, 2009. -   U.S. Pat. No. 7,565,742 Methods of Designing Lobe-Type Rotors. Jul.     28, 2009 -   U.S. Pat. No. 7,563,741 Methods of Designing Lobe-Type Rotors. Jul.     28, 2009. -   U.S. Pat. No. 7,562,450 Methods for Designing Single-Lobe Rotors.     Jul. 21, 2009. -   U.S. Pat. No. 6,776,594 Rotor Mechanism. Aug. 17, 2004. -   U.S. Pat. No. 5,674,051 Positive Displacement Pump having     Synchronously Rotated Non-Circular Rotors. Oct. 7, 1997. -   U.S. Pat. No. 5,660,535 Method of Operating a Claw-type Vacuum Pump     and a Claw-type Vacuum Pump Suitable for Carrying Out the Method.     Aug. 26, 1997. -   U.S. Pat. No. 5,401,151 Vacuum Pumps. Mar. 28, 1995. -   U.S. Pat. No. 4,324,538 Rotary Positive Displacement Machine with     Specific Lobed Rotor Profiles. Apr. 13, 1982. -   US 2014/0227122 Claw Pump. Aug. 14, 2014. -   The London Gazette Nov. 12, 1869.

BACKGROUND TO THE INVENTION

Rotary Claw pumps have been in common use for many years. They are a variant of the Roots pump which is named after the American inventors and brothers Philander and Francis Marion Roots, founders of the Roots Blower Company, Connersville, Indiana, who first patented the basic design in 1860 as an air pump for use in blast furnaces and other industrial applications, London Gazette Nov. 12, 1869. The Roots pump is a machine for compressing or evacuating air or gas by the rotation of a meshing pair of lobed rotors in a closely fitting case. Each rotor consists of a planar lobed or dumbbell shape extruded vertically from the plane with arrangements to allow for their rotation. The case consists of a peripheral enclosure which closely encompasses the rotor diameters and an inlet plate and an outlet plate that closely fit the respective flat faces of the rotors. Rotor pairs are arranged to rotate in opposite directions and their rotation is synchronized.

Rotary claw pumps have one or more pairs of intermeshing rotors which are fashioned into claw shapes. The claws of these devices are arranged so that the profile of one claw relative to the center of rotation of its rotor is substantially the conjugate of the other so that the claws mesh without interference but with tight clearances between the rotors.

Tight clearances are required between the paired claws to minimize the flow of fluid between the points of closest approach of one claw with corresponding part of the body of its pair, between the tips of the said claws and the casing and between the flat face of the rotors and the casing. The tighter the clearances the better within the constraints imposed by the sums of all the manufacturing and assembly tolerances.

This type of design is illustrated in U.S. Pat. No. 4,324,538.

To date, marketed rotary claw pumps using rotor pairs with two or more claws per rotor, although having claws of substantially conjugate shapes, have claws of different profiles; for example, see FIG. 1, items 1 and 2.

Definitions

For the purposes of this patent the following definitions apply:

Viewed from a plane perpendicular to the axis of rotation, Angle Theta (0) is the smallest angle between the line joining the centers of rotation of a synchronized intermeshing pair of clawed rotors and a rotation of this line about the center of rotation of one of the said rotors when the rotors turn from the position where the tips of their claws are at the point of their closest approach to the position where the tips of their claws are at their furthest apart. This is illustrated in FIG. 3 by rotors 1 and 2.

DESCRIPTION OF EMBODIMENTS

It is advantageous to have claws of the same nominal profile as this provides economies of scale during manufacture, lower inventory costs and greater ease of replacement with less chance of error. This is especially so when rotary claw pumps of multiple stages are used and when the claws are molded or extruded rather than machined as this greatly simplifies the production process and significantly lowers costs. In the case of multiple stage pumps it is possible to arrange that the thickness of each successive rotor pair does not vary in profile but has a thickness approximately equal to the thickness of the first stage rotor pair multiplied by the inverse products of the compression ratios of the preceding stages. By this means the overall compression ratio of the device is shared approximately equally between the stages and this allows the maximum compression ratio to be achieved from a given number of stages from the minimum physical size pump. Such change of thickness is easily accommodated during molding and extruding processes without necessarily having to build separate individual dies. Molding and/or extruding of parts also allows the use of different molding materials utilizing the same mold thus providing for changes in the fluid being pumped and changes in design pressures and temperatures.

For example, for pressures up to 200 psi and temperatures up to 200 C such as encountered in air compressors, PolyPhenylene Sulfide might be chosen, while for somewhat lower pressures and temperatures up to 100 C, Glass filled Polypropylene might be chosen, while for temperatures down to −90 C such as those encountered in low temperature refrigeration, Nylon might be used. Where multiple stage compression is used all claws may be of the same nominal profile differing only in their thicknesses as appropriate to the volume throughput. Such thickness changes are easily accommodated in molding and extruding processes.

Thus the use of molded claws with the same nominal profile greatly simplifies the manufacturing process, lowers costs and increases flexibility.

The use of claws with the same nominal profile also ensures a more even rate of compression of the fluid with rotation and eliminates the possibility of over compression as described in U.S. patent application Ser. No. 14/347,206 publication no US 2014/0227122.

It is advantageous for the claw profile to change uniformly. If the change in profile is not uniform, noise is generated and mechanical fatigue can result in failures.

Various methods have been proposed to calculate claw profiles, see U.S. Pat. Nos. 7,594,323, 7,565,742, 7,565,741, 7,562,450 and 6,776,594, but none of them result in claw profiles which are nominally the same. The prior art regarding intermeshing rotors even with single claws as in, for example, U.S. Pat. Nos. 8,517,701, 8,308,458, 5,674,051 and 5,401,151, while it shows claws that are relatively close in shape, nonetheless, does not disclose claws which have the same nominal profile.

Conjugate claw profiles are achieved by ensuring that for the convex parts of the claw profile, the rate of increase of root diameter of one claw per unit change in angle of rotation is nominally matched by the rate of decrease of the outer diameter of the conjugate claw over the same change in angle of rotation. However, this condition while necessary to achieve lower noise and fatigue, is not a sufficient condition. Most such profiles are arranged so that there are no slope discontinuities in the profile, for example, as in FIG. 1.

U.S. Pat. No. 8,887,593 does describe claws of the same profile using a combination of epicycloids and arcs, but the rates of decrease and increase of the profiles with change in rotation angle are not uniform. Thus, although there are no slope discontinuities at the junctions of the curves there are slope discontinuities between the different curves, thus causing pulsations and noise.

In this invention, to ensure that the claw profiles are nominally the same, the aforesaid rates of increase and decrease with change in rotation angle are made uniform as in FIG. 2.

This ensures an even compression rate without causing any pulsations, and the profile is simply calculable without the need for any curve equations and is easily transcribed to CNC machines for manufacturing molds or parts.

Some characteristics of rotary claw pumps are:

1. There is always a percentage of fluid which is recycled due to the fluid trapped between the claws; for example, see FIG. 2 hatched region 2.3 which depicts a two claw per rotor design (items 2.1 and 2.2). When the ratio of the Claw Radius to half the inter-Claw pitch is 1.333 where the claws of the same nominal profile are used (designated Case A), 15.72% of the input flow is recycled. 2. They have a limited practical compression ratio for compressible gases due to gas bypass past the mating parts of the claw and also past the claw tips and claw sides and the casing. The maximum practically allowable compression ratio is usually taken as being about 3:1. 3. They are not generally tolerant of substantially incompressible fluids. 4. They have very little wear due to the lack of rotor to rotor and rotor to wall contact. This makes them very suitable for air conveying of dust or particle streams, for example, in cotton gins. 5. They require no lubrication in the claw chambers. Thus they produce oil free compressed fluids. This is especially important in food processing and medical uses. 6. There are no inlet or outlet valves. 7. It is possible to gang rotor pairs in series on their respective shafts so that multiple stage compression can easily be achieved.

It is possible to make rotary claw pumps with one to many claws per rotor. The percentage of recycled flow is less as the number of claws per rotor decreases. A rotor with a single claw, sometimes called a Northey rotor is, however, inherently unbalanced and is only suitable for very low revolutions per minute. A rotor with two claws is inherently balanced and gives rise to the lowest recycled flow percentage for inherently balanced rotors regardless of the number of claws per rotor. For this reason, although the concepts presented have application to rotors with any number of claws per rotor, the description given is described using two claws per rotor designs.

In the case of 1 above, when the fluid is compressible, the recycled fluid is at a greater pressure than the inlet pressure to the pump and the fluid has been heated by essentially adiabatic compression.

The maximum compression of the recycled fluid occurs when the rotors are in the position shown in FIG. 2 and at this point the hatched region 2.3 in FIG. 2 is about to be discharged back to the inlet as the rotors rotate further.

In the case of compressible gases the compression ratio at this point is 1.2515 to 1, that is, the volume has been reduced to 0.799 of the start volume. This recycled fluid is released substantially isentropically back into the inlet and expands, thus having the effect of causing pulsation and very slightly lowering the intake volume per revolution (typically by 0.06%). The work done in compressing this recycled fluid is not useful work as far as delivering the fluid at its output pressure and volume is concerned and is essentially wasted.

It is advantageous when pumping to minimize pulsation in the pump. This may be done by ensuring that when the fluid is compressed to the same pressure as the outlet or designed outlet pressure, the part of the chamber where the fluid is under compression is opened to the outlet.

This is designated Mode 1.

Rotary claw pumps have in the past only been used for compressible gases (albeit in some cases gases carrying some solids) but not for incompressible fluids. They have also not been used for refrigeration duty pumping condensable refrigerants. This lack of use has primarily been due to serious overpressure situations caused by pumping liquids or even pumping two phase liquid gas mixtures which can lead to catastrophic damage.

To make rotary claw pumps suitable for use in the case of a substantially incompressible fluid, the latest position in the rotation where the fluid to be discharged must be first be connected to the outlet is when the tips of rotor claws adjacent to the fluid to be discharged are at their point of closest approach as that is when compression of the fluid would otherwise commence. From that point on in the rotation, the fluid must be continuously connected to the outlet during all parts of the subsequent rotation until the said claw tips are at their point of being furthest apart as that is when the fluid has been discharged. This is illustrated in FIG. 3. Hatched region 2.3 is a region that is uncovered by rotor 2.2 as it rotates and this region is connected to the outlet. Region 2.3 is the maximum sized region that fulfils these requirements while beginning to open at the latest position described above and giving rise to the lowest exit velocity and exit losses. It is of course possible to open such a region earlier in the rotation, but it must not so be early that it causes a direct connection between the inlet and outlet. Opening such a region early does not reduce discharge losses as no discharge takes place at that time. It is, however, beneficial to open the said region a little earlier than necessary in order to ensure that no fluid pulsations occur.

An additional and alternative method of relieving the pressure on the fluid is by providing one or more output valves such as FIG. 3 item 3.8 which opens when pressure builds up in the region containing the fluid to be discharged. It is preferable that said valves discharge to either the outlet or the inlet of the stage they are relieving rather than discharging externally, and it is preferable to build in such valves internally into the pump. Valves such as this may be placed wherever necessary.

When used in addition to the method of port uncovering by rotor rotation, this serves as a safety valve. When a mixture of compressible and incompressible fluids is the working medium, such as is the case in many refrigeration uses, then the outlet may be configured to open at a location between the claw positions shown in FIGS. 3 and 4. In this situation the presence of some compressible fluid allows compression to take place while there is no connection to the outlet without causing a possibly damaging overpressure situation. In order to avoid pulsations it is still desirable to configure the outlet opening to follow the requirements of Mode 1. To prevent any possible overpressure, especially due to any abnormal increase in the percentage of incompressible fluid in the mixture, it is highly desirable to incorporate a safety valve as in FIG. 3 item 3.8. Thus in this situation, it is not necessary to operate the Pump without the compression described in U.S. Pat. No. 5,660,535.

FIG. 3 shows the claws 3.1 and 3.2 in the position where compression starts, and hatched region 3.3 shows the maximum allowable outlet opening that is uncovered as rotor 3.2 rotates.

FIG. 4 shows the claws 4.1 and 4.2 in a position where, for a compressible fluid, the fluid in FIG. 4 hatched region 4.5 might equal the outlet pressure and FIG. 4 hatched region 4.3 shows the maximum allowable outlet opening to be uncovered.

One of the major issues in refrigeration duty is the necessity for compressor lubricants. These are required for all types of pumps including screw compressors. A large industry is based on providing suitable lubricants for the multitude of refrigeration gases and temperature pressure conditions that apply. It is very important that refrigeration lubricants be recovered before the refrigerant expands in the evaporator as otherwise waxing or coating of the evaporator surfaces results in significant heat transfer reduction. This recovery can be quite difficult to achieve: normally demisters and other similar devices are used for this purpose and, again, a large industry is based on supplying these. The lubricants also need periodic replacement.

In the case described above, it is possible to use a claw rotor pump without any lubricant in refrigeration systems, and this gives rise to a major simplification in such systems and an increase in longevity and reliability. Refrigeration cycle systems use a condenser after compression to cool the refrigerant. In many cases the required compression ratio is beyond the maximum recommended limit of 3:1 for claw rotor pumps, but this gives rise to a further advantage in that two or more stages of compression are then used with intercooling between stages. As discussed below, this gives rise to a significant reduction in the power requirements due to the fact that the multi-stage rotary claw pump performs much closer to an isothermal compressor than a single stage adiabatic compressor. It is also possible to arrange rotary claw pumps so that they are semi-hermetic by coupling them by sealed means to a hermetically sealed motor in a similar fashion to that presently done with screw pumps. When utilizing hermetically sealed motors in this fashion it is also advantageous to use motors that do not require the passage of refrigerant through or around the motor in order to cool the motor. Any heat introduced into the refrigerant by cooling the motor must be subsequently removed by a condenser and thus adds to the size and cost of such a condenser.

When compressing a compressible fluid or a two phase liquid gas fluid in normal rotary claw pumps which are configured in Mode 1, when the rotors rotate past the position shown in FIG. 3, the point where the tips of rotors 3.1 and 3.2 are closest together, the tip of one claw (FIG. 3 item 3.7) begins to compress the fluid in FIG. 3 hatched region 3.5 as at this stage the exhaust is shut off. For compressible fluids, a solution to problem of the energy wasted by recycled fluid may be provided by, as the rotors turn past the position shown in FIG. 3, opening a region (FIG. 3 hatched region 3.4) to the inlet. The maximum size of this region uncovered by claw 3.1 as it rotates is given by area inside the intersections of the line between the centers of the rotors, a rotation of the line between the centers of the rotors by the angle Theta (0) about the center of Rotor 3.1, the concave curve of rotor 3.1 and the tangential extensions from the ends of said curve when the tips of rotors 3.1 and 3.2 are at their closest together, the minimum rotor diameter of rotor 3.1 and the convex curve of rotor 3.1, and the tangential extension from the closest end of end of said curve to the center of rotor 3.1 when the tips of rotors 3.1 and 3.2 are at their furthest apart as in FIG. 2. Hatched region 3.4 begins to be uncovered when the rotors turn from where their tips are at the closest point of approach and remains at least partially open until these same rotor tips are at their furthest apart. The tip of claw 3.2 forces surplus gas from region 3.5 back to the inlet, thus greatly reducing the compression in that region. While the above description aids in constructing region 3.4, a more general definition is that after the rotors continue to turn from where their rotor tips are at their closest point of approach and until the said rotor tips are at their furthest apart, a region connected with the tips of both rotors 3.1 and 3.2 is at least partially connected with the inlet and fluid is allowed to pass from said region to the inlet.

The opening of the said region to the inlet is important as the recycled gases are maintained at close to the inlet pressure and temperature thus eliminating pulsation from this cause as there is little or no isentropic expansion of the recycled gases, and thus no input energy is required other than that due to frictional losses. In the example given, the pressure in this region is reduced to close to the inlet pressure and the compression ratio is reduced from 1.2515 to close to 1. The actual percentage of energy thereby saved depends on the output compression ratio and the output pressure. In Case A, for example, although 15.72% of the flow is recycled, it is recycled from a compression ratio of 1.2515, which is normally less than the designed compression ratio of the pump. Therefore, less than 15.72% of the total energy is wasted. If the designed output compression ratio was 3:1 then, very approximately, the amount of wasted recycled energy would be 1.2515/3*15.72%=6.55%, and, if the designed compression ratio were 2:1, then the wasted recycled energy would be approximately 9.83%.

The saving of this amount of energy is useful but there is a penalty. The overall output volumetric rate of the pump is reduced for a given physical size pump and, therefore, the pump must be increased in size in order to achieve a particular designed output. In Case A this means that the pump throughput must be increased by 15.4%. This can be done by increasing all dimensions by 4.87%, the claw thickness by 15.4%, or the diameter by 7.4%.

In the case of substantially incompressible fluids it is also advantageous to minimize pulsation caused by passage of compressed fluid through FIG. 3 passage 3.6 by opening the said region to the outlet rather than the inlet to avoid reducing the volumetric rate of the pump.

Rotary Claw Pumps are usually directly coupled to electric motors and the general layout of these is that the output shaft of the motor is attached to a coupling/sleeve which itself is attached to one of the pump rotor shafts. The next items on the pump shaft are the gears and a bearing which are interchangeable as far as their order is concerned. The gears are usually enclosed by a lubrication chamber into which the bearing is sometimes included. These are followed by an intake casing which may in some cases partially enclose the bearing and/or the gearbox enclosure.

The next item on the pump shaft is the first pair of rotors. An illustration of this is shown in U.S. Pat. No. 8,517,701 FIG. 1. When the motor is a “C Face” unit there is a method of attachment, and location is provided between the intake casing and the motor.

It is advantageous to move the gearbox to beyond the furthest pair of rotors from the output shaft of the driving motor and eliminate the said bearing and, where possible, use the nearest rotor to the motor output shaft as the coupling/sleeve. This not only eliminates a bearing and often a sleeve or coupling but it also eliminates misalignment between the pump and the motor and also prevents such misalignment from causing shaft and bearing failures. The length of the pump is also shortened and cost is reduced. It is also preferable to ensure that the pump may be removed from the motor without dismantling of the pump itself.

When the ratio of maximum rotor radius is R and the pitch between the centers of a pair of rotors is Rp as in FIG. 3, then C is defined as R/Rp. The root radius of the rotor is then 2*Rp−R. For validity C must be greater than 1 and less than 2. When C is 1 no pumping takes place as R=Rp and there is no interlocking of claws. When C is 2 the root diameter of the claw is zero and the claw ceases to exist and there is no room for a shaft or any thickness of claw at the shaft diameter. K is defined as the ratio of the thickness of the thickest claw pairs to the radius R. Thus the thickness of said claw pair is K*R and is therefore proportional to K where K is greater than zero. The fluid intake volume and, therefore, for a given Radius R the output mass of fluid produced by a Rotary Claw Pump per revolution is proportional to K*(1/C−1/C²) and for a given K is proportional to (1/C−1/C²). This function reaches a maximum at 2; hence, the closer C can come to 2, the more output mass per revolution. In practice, space must be allowed for a drive shaft and a minimum claw thickness around the shaft, and this means that C may be considered conservative when set to 1.33, aggressive when set to 1.5 and too large when set above 1.66.

The frictional losses due to the passage of fluid past the rotor are proportional to surface area where the flow takes place and, for a given radius R, are proportional to 1+K. The fluid intake volume and therefore the output mass of fluid produced by a Rotary Claw Pump per revolution is also proportional to the claw thickness K*R. Therefore, the frictional losses per unit of output mass produced per revolution for a given claw radius are proportional to the function (1+K)/K. This function decreases with increasing K and suggests that large K values should be used. However, at K values above approximately 5, significant axial flow is required in the rotors, and this can lead to cavitation in the case of substantially incompressible fluids and over pressure during the forcing of the fluid to discharge, therefore it is suggested that a K ratio of 5 should be regarded as a practical limit.

As described above, the shafts on Rotary Claw Pumps are normally coupled together by gears which are invariably placed adjacent to the output shaft of the driving motor. Such gears usually require lubrication and, since one of the features of Rotary Claw Pumps is the ability to avoid contamination of the pumped fluid by lubricants, it is important to ensure that no lubricant can contaminate the fluid. This may be done whether the pump is used as a suction device or for the delivery of fluid by placing the gears and their containment further away from the output shaft of the driving motor than the furthest pair of rotors as described above. In this position it is possible to seal and contain the gears so when the pump is a fluid delivery device the pressure of the fluid, being at or above the pressure in the lubricant, prevents any lubricant from entering the fluid stream. When the pump is a vacuum device it is not normally a requirement to avoid lubricant leakage into the discharged fluid, but if that happens to be a requirement, the provision of a passageway connected to atmosphere between the gearbox and the pump suffices to ensure that any lubricant leakage from the gearbox is discharged separately from the pump discharge.

The gearbox may be positioned so that it is at least partially within the outlet chamber for the pump. Such positioning shortens the pump and lowers the cost and also allows the gearbox to be serviced without removing the pump from the motor.

It is important to minimize pulsation in multistage Rotary Claw Pumps. Usually, successive rotors have their tips aligned as shown for example in U.S. Pat. No. 5,401,151 FIG. 1. As is discussed in U.S. Pat. No. 8,517,701, this gives rise to inter-stage pressure pulsations and also may limit the overall compression ratio that can be achieved. U.S. Pat. No. 8,517,701 prescribes a method of reducing this pulsation. A more general method is to arrange the phasing of successive stages so that each stage is taking in fluid when the preceding stage is discharging fluid. It is preferable that the said intake takes place over the entire period the previous stage is discharging, but even if the overlap is not total, a significant reduction in pulsation occurs.

There are some circumstances in which both suction and compression and/or fluid discharge is required. It is possible to arrange Rotary Claw Pumps to provide both from the one pump by arranging compression/discharge and suction stages in succession on the same shafts thus avoiding the need for two or more separate pumps. Thus, multiple levels of compression/discharge and/or vacuum may be provided from one pump.

Compressors are usually provided with intake filters and sometimes are provided with means for removing water from the compressed fluid. The intake filters are intended to increase longevity of the compressors but they require periodic servicing. At present, the life of screw compressors is of the order of five years. The intake fluid is often air, and in the following description this is used although the description is also applicable to other fluids.

It is advantageous to clean the intake fluid by scrubbing the air in a water droplet stream as this method requires no periodic maintenance. Water mist remaining in the air may be removed by a traditional demister obeying Souder-Brown design criteria. This leaves the air saturated with water vapor. This vapor may be substantially removed by chilling the air. This may be done by a single heat exchanger when chilling down to approximately 4 degrees C., and lower temperatures may be achieved by using heat exchangers in parallel where one exchanger is chilling the air while another is defrosting. Following the chilling, water mist remaining may be removed by another demister. To complete the air cleaning process an electrostatic precipitator may then be used.

This means provides air that is substantially free of particulates, has a known and low level of water vapor, and dissolvable gases have been substantially removed. This gives rise to longer compressor life. However, there is also an improvement in the volumetric efficiency of the compressor due to the lowering of the intake temperature. When the ambient temperature is 25 C and chilling is done to 4 C, this gain is 7.5%. Further, there is a substantial reduction in energy requirement caused by not compressing the unwanted water vapor and dissolvable gases, and this is achieved without the need for periodically replaceable elements. Thus, the method of cleaning the intake air to the compressors by water scrubbing followed by demisting, followed by chilling, followed by demisting, optionally followed by electrostatic precipitation is very advantageous. It is particularly so when the air is to be used for separation equipment such as in pressure swing absorption apparatus. The molecular sieves used in such equipment benefit not only from the air being clean but also are more effective and have greater longevity as they have much less water vapor and dissolvable gases to remove.

In the matter of item 2 in the above list this limited compression ratio can be an advantage. Unlike other classes of compressors, with the rotary claw pump it is a simple matter to arrange for successive stages of compression and intercooling by mounting compressor stages on the same drive shafts. This is also a very low cost configuration.

Compression of gases is essentially adiabatic, giving rise to significant amounts of energy going into heating the gas. This heat is normally lost and represents inefficiency in the process.

Ideally, compression would be carried out isothermally. FIG. 5 shows the difference between isothermal and adiabatic compression using Nitrogen gas being compressed from 1 atmosphere absolute to 9.5 atma, a typical range for commercial air compressors and an overall compression ratio of 4.94. The relative energies required for adiabatic compression and isothermal compression are 3.179 and 2.252, respectively. It should be noted that while the mass of compressed gas produced by both processes is the same, the volume produced adiabatically is greater than that produced isothermally, but it is at a higher temperature. When that gas cools, it reverts to the same compressed volume as in the isothermal case. Thus, 29.2% of the energy put into compressing adiabatically is lost.

The required overall compression ratio is beyond the recommended limits for rotary claw pumps; hence such a compressor would use two or more stages. This allows for intercooling between the stages and this reduces the energy requirements and approaches the isothermal energy more and more closely as stages are added. FIGS. 5 to 9 show Pressure Volume graphs with Temperature shown on the secondary Y axis. The work done in compression is the integral of the PV curves. In the case of a two stage design with compression ratios of 2.2 (See FIG. 6), only 15.1% of the energy put into compressing adiabatically is lost, and in the case of a three stage design with compression ratios of 1.7 (see FIG. 7), only 10.5%. Thus, an intercooled multistage rotary claw pump approaches isothermal efficiencies quite closely and is much more efficient than single stage adiabatic compression.

A particular use where it is very important to maximize efficiency of compression is in the case of storage and regeneration of electric power. Presently, movement towards the replacement of fossil fueled power generation with renewable generation is being severely limited by the lack of suitable energy storage means. The problem is that the availability of renewable power from sources such as wind and solar does not correspond with the load requirement of electric power, and the magnitude of the renewable sources and the lack of storage is now such that base load electric power stations are forced to vary their outputs very rapidly. This is in most cases not possible, so it is leading to power dumping, curtailment and paying users to take electric power.

Various methods for storage and regeneration of power have been proposed such as pumped Hydro, battery, storage of rotational energy, and pumped air systems. Each of these has severe limitations such as availability, cost, longevity, inefficiency, safety, footprint and limited storage.

Hydro storage can be up to 80% efficient but it is not generally available. Batteries are expensive, have limited life and have safety issues.

Compressed gas systems have suffered from inefficiency and cost problems. Attempts have been made to overcome the inefficiency issue by approaching isothermal compression by compressing very slowly, but this in turn requires very large heat exchange systems which are costly. Another method that has been attempted is to introduce a heat transfer fluid, usually water, into the compressed gas, usually air, and store and recover the heat energy, but various problems have arisen with this approach.

A multi-stage rotary claw pump with low compression ratios per stage can be used to overcome the issues of cost and efficiency. FIG. 8 shows a twelve stage compression taking air from normal atmospheric pressure up to 3000 psi with intercooling between each stage. This can easily be accomplished by using an electric motor with dual output shafts and by applying 6 stages of one diameter to one shaft and six stages of a different diameter to the other shaft. As has been previously shown, it is possible for such a system to approach the efficiency of isothermal compression. A further improvement in efficiency is possible by using air to fluid counterflow heat exchangers between each compression stage and by storing the energy removed in insulated storage tanks. It is preferable to use a system that transfers the heat exchange fluid between an input tank which is at the lower of ground water and ambient temperatures but above the freezing point of the heat exchange fluid, and an output tank which receives the heated fluid. In some situations it is possible to further increase efficiency by providing means to change the temperature of the input tank as ambient temperatures vary from day time to night time temperatures.

By this means as shown in FIG. 9, the temperature of the compressed gas may be cooled below the isothermal temperature at each intercooling stage as opposed to merely being cooled to a temperature slightly above the isothermal temperature as shown in the previous examples and as is the case with parallel flow heat exchangers or fluid injection systems such as described in U.S. Pat. No. 9,444,378. This increases the efficiency of the system such that the compression losses are in the range 5 to 7.5% of the input energy. Rotary claw pumps are inherently able to run in the reverse direction as motors and the drive motor can also function as a generator. Because of this, the stored energy may be converted back into electric power, and the stored energy in the fluid may be recovered by using the same counterflow intercooling heat exchangers as interheating heat exchangers to reheat the expanding air to slightly above the isothermal line at each stage. The use of counterflow as opposed to parallel flow heat exchangers, which is the defacto case with fluid injection systems such as described in U.S. Pat. No. 9,444,378, is important as counterflow heat exchange increases the efficiency of the system as described above.

It will be appreciated that when an individual compressor changes from receiving energy to delivering energy or vice versa the direction of rotation changes. This imposes a delay when switching directions of energy flow due to the need for the rotors to slow down, stop and speed up in the reverse direction. Such a delay may be avoided in systems which have multiple compressors by selectively arranging for some compressors to stop or switch direction while others continue as they were.

Multiple stage compression systems typically use the same compression ratio for each stage. This does not necessarily result in the lowest losses and highest efficiency. With rotary claw pumps it is possible to vary the compression ratios of each stage quite simply by changing the thicknesses of rotor pairs. Thus, it is possible to optimize the overall efficiency of the system by choosing appropriate compression ratios, which may or may not be the same, for each of the stages. By using these means the efficiency of electric power to compressed air to electric power systems can be in the range 85 to 90%.

Compressed air storage cylinders are readily available at pressures up to and exceeding 3000 psi. It is important that the coming fluid, usually air, is clean and dry to avoid accumulation of unwanted material within the storage vessels, to avoid corrosion, and to avoid waste of energy compressing unwanted vapors. Because of this, it is useful to use the previously mentioned cleaning arrangements upstream of the compressor(s) and also to cool the delivered compressed fluid stream preferably to close to the freezing point of water and to precipitate out condensed vapor so as to ensure that there is no condensation of water when the storage is exposed to normal ambient temperatures. 

1. A Rotary Claw Pump wherein at least one pair of intermeshing rotors has the same shape within a tolerance of plus or minus 3 percent of the radii from the edges of their shapes to their centers of rotation and wherein on a convex portion of the claws there is a nominal uniform rate of change of radius with angle of rotation.
 2. A Rotary Claw Pump as in claim 1 where there are multiple stages wherein the thickness of each successive rotor pair does not vary in profile but has a thickness within 20% of the thickness of the first stage rotor pair multiplied by the inverse products of the compression ratios of the preceding stages.
 3. A Rotary Claw Pump wherein a discharge port is at least partially uncovered at any position of rotation from where a pair of intermeshing claw rotor tips are at their closest point of approach to where the said claw tips are at their furthest apart and wherein said discharge port is at least partially connected with a region connected with the rotor tips of both the said claws during said positions of rotation and the discharge port thereby at least limits pressure increase in the said region.
 4. A Rotary Claw Pump as in claim 3 wherein the fluid being pumped is at least partly compressible and wherein the said discharge port is connected to an inlet region for the said rotors.
 5. A Rotary Claw Pump as in claim 3 wherein the fluid being pumped is substantially incompressible and wherein the said discharge port is connected to an outlet region for the said rotors.
 6. A Rotary Claw Pump wherein the working fluid is substantially incompressible and wherein at any position of rotation from where a pair of intermeshing claw rotor tips adjacent to the fluid to be discharged are at their point of closest approach to where the said claw tips are at their position of rotation when they are furthest apart said fluid is at least partially discharged.
 7. A Rotary Claw Pump wherein the working fluid has a liquid and a gas phase and/or the working fluid may be a refrigerant and wherein a discharge passage for said fluid is at least partially opened when a pair of intermeshing claw rotors rotate from a rotational position of the tips of said claws where the pressure in said fluid is within plus or minus 15% of a designated output pressure and wherein said discharge passage remains at least partially opened until the said claw tips are at their furthest apart.
 8. A Rotary Claw Pump wherein over pressurizing is minimized by the provision of one or more valves to relieve one or more internal regions of said pump.
 9. A Rotary Claw Pump wherein the nearest rotor to the driving motor shaft has no intervening bearing between it and the shaft of the driving motor.
 10. A Rotary Claw Pump as in claim 10 wherein one of the first pair of rotors adjacent to the driving motor is used to couple the motor shaft to the pump.
 11. A Rotary Claw Pump wherein the gears coupling the driving shafts are further way from the output shaft of the driving motor than the furthest pair of rotors from the output shaft of the driving motor.
 12. A multi-stage Rotary Claw Pump wherein the phasing of successive stages is arranged so that each stage is taking in fluid during at least 80% of the time when the immediately preceding stage is discharging fluid.
 13. A Rotary Claw Pump wherein various stages of the pump are arranged to separately compress/discharge and/or provide suction.
 14. A Compressor wherein at least part of the intake gas is cleaned by scrubbing with water, followed by demisting, followed by chilling, followed by demisting and optionally followed by electrostatic precipitation.
 15. A Compressor as in claim 14 wherein the gas is air.
 16. A Compressor as in claim 14 wherein the compressor is of the Rotary Claw type.
 17. A Compressor as in claim 14 wherein the gas is used as the input to a pressure swing separator.
 18. A multi-stage Rotary Claw Pump system to retrievably store electric power as a compressed gas wherein at least two stages of compression have counterflow intercooling heat exchange means arranged to cool the compressed gas after compression and heat the decompressed gas after expansion and wherein, during compression, the heat exchange fluid used in at least some of the said counterflow heat exchangers is drawn from a first fluid store, is passed through said heat exchange means, and is sent to a second fluid store.
 19. A system as in claim 18 wherein during expansion the heat exchange fluid used in at least some of the said counterflow heat exchangers is drawn from a fluid store which may be said second fluid store, is passed through said heat exchange means, and is sent to another fluid store which may be said first fluid store.
 20. A system as in claim 18 wherein the temperature of fluid in said first fluid store approaches the lower of ground water and ambient temperatures, but is above the freezing point of the heat exchange fluid and/or wherein the temperature of fluid in said second fluid store is at or above ambient temperature.
 21. A system as in claim 18 wherein the compression ratios of stages are at least partially optimized to reduce the loss of energy during compression.
 22. A system as in claim 14 wherein compressed fluid or fluid to be compressed is optionally cooled and condensate is precipitated and optionally removed. 